This relates to a method and apparatus for controlling a pipeline for the transmission of gas. More particularly, it relates to a method and apparatus for controlling the operation of compressors commonly used in pressure boosting stations along the pipeline.
Pressure boosting stations are used to maintain a desired discharge pressure, suction pressure and/or flow rate of the gas in the pipeline. Generally, pipeline boosting stations require compressors that have high capacities and low compression ratios (discharge pressure to suction pressure), typically, a ratio on the order of 1.3 to 1.
Centrifugal compressors are normally used to provide such high capacities and low compression ratios. Such compressors have lower maintenance costs, lower installation costs in the large capacity sizes, and greater reliability of service than reciprocating compressors. However, the centrifugal compressor is designed to operate at a particular point; and, as shown in FIG. 1, the efficiency of the compressor drops off rapidly if conditions deviate widely from the design point. FIG. 1 is a plot of the efficiency of a compressor versus inlet flow and head. The compressor operates most efficiently along the curve labeled optimum efficiency. The compressor may be operated at lesser levels of efficiency to the right or the left of the optimum efficiency line if it is not possible to achieve operation at optimum levels. Operation to the left of the optimum efficiency line emphasizes the production of more head than is optimally necessary, while operation to the right of the optimum efficiency line emphasizes the production of more flow than is optimally necessary. Thus, for a given compressor speed the compressor can be operated so as to produce any head and flow between the surge and stonewall lines at the efficiency level associated with the head and flow desired. In the prior art, when the operating speed of a compressor is changed from one level to another there is little or no control exercised over how that change is to be affected. As a result, in the prior art, there may be significant losses in operating efficiency when it becomes necessary to change the operating point of the compressor.
A further problem in the use of centifugal compressors is that their operating range is limited by "stall" and "surge" conditions. Stall occurs in a centrifugal compressor when the flow rate is increased to a point where the flow reaches a maximum rate for the given inlet conditions and speed of the compressor. For a given suction pressure and speed, discharge pressure decreases approximately linearly with the flow rate through the compressor. When the compressor stall point is reached, any further attempt to increase flow causes a rapid decrease in discharge pressure; thus, a limit is placed on the capacity of the compressor. Since the compressor is usually operated at varying speeds, stall conditions occur along a line known as the stall line (sometimes called "stonewall line") as shown in FIG. 1. When the compressor is operated at the stall line, no further increase in flow can be obtained for the particular set of conditions.
Surge occurs in a centrifugal compressor when the flow rate through the compressor is decreased to a point that is insufficient to maintain discharge pressure at a higher level than the line pressure into which the compressor is discharging. The discharge pressure falls below the line pressure momentarily, and a sudden reversal of flow occurs through the compressor. This sudden reversal of flow causes the compressor discharge pressure to rise and line pressures to drop until discharge pressure rises slightly above line pressure again. Since discharge pressure is now above line pressure, flow makes a second reversal and resumes its original direction. However, as soon as flow is resumed out of the compressor, discharge pressure begins to drop and line pressure rises until line pressure is again higher than the discharge pressure. The result is that the flow oscillates and shocks are transmitted to the compressor which vary from an audible rattle to a violent shock that can damage the compressor impeller, and even possibly bend the compressor shaft.
The points at which surge occurs are shown plotted as the surge line in FIG. 1. The compressor should not be operated at or to the left of the surge line, and since the compressor cannot be operated to the right of the stall line, the operating range of the centrifugal compressor is between the surge and the stall lines. For a given discharge pressure, the centrifugal compressor can be seen to have a narrow operating range. For this reason, the centrifugal compressor can only be used where conditions do not vary widely. Because the flow rate through the pipeline does tend to vary widely, it is common practice to locate several compressors at each boosting station and to vary the number of compressors in use so as to maintain the desired operating parameters.
The equation for the surge line is: ##EQU1## where P.sub.D =Discharge Pressure
P.sub.S =Suction Pressure PA1 K.sub.1 =Constant PA1 T.sub.S =Suction Temperature PA1 Q.sub.A =Volumetric flow measured at inlet conditions PA1 T.sub.B =Base Temperature
The relation between the actual inlet flow, Q.sub.A, and the inlet flow at base conditions, Q.sub.I, is given by ##EQU2## where P.sub.B =Base Pressure
The inlet flow at base conditions can also be shown to be ##EQU3## where h=differential pressure across an inlet orifice plate
Substituting, equation (3) into equation (2) and rearranging ##EQU4## where K.sub.2 is a constant. Substituting equation (4) in equation (1) ##EQU5## where K.sub.3 is a constant.
Hence, at any point along the surge line, the pressure difference across a centrifugal compressor is proportional to the pressure differential across the inlet orifice plate; and by measuring the pressure difference across a centrifugal compressor, a minimum inlet orifice differential can be calculated which will prevent the compressor from going into surge. In the prior art, inlet flow and hence the pressure differential across an inlet orifice is maintained at a minimum level by apparatus which senses the compression ratio of the compressor and, when it approaches surge conditions, bypasses some of the discharge gas through a pneumatically operated bypass valve back into the suction side of the compressor. This maintains a minimum throughput through the compressor even when the minimum system throughput is less than the level which would put the compressor into surge if there were no feedback of gas through the bypass valve.
This technique, however, is wasteful of energy and the energy is lost in raising the temperature of the gas. If the temperature gets too high, it is necessary to shut down the compressor which is likely to affect the operation of the other compressors at the same boosting station and perhaps the operation of the entire pipeline. Moreover, this prior art technique provides no means of eliminating the surge condition once it has begun. As a result, compressors that fall into surge condition can remain operating with their bypass valves open for long periods of time (days or weeks)if they do not overheat. Moreover, once the surge condition is noticed by the pipeline operator, it is usually necessary to shut the compressor down and modify the pipeline operating parameters before the compressor can be used again. As will be apparent, coping with surge conditions with such prior art techniques can seriously affect the supply of gas and/or the operating efficiency of the pipeline.